The present invention relates to a vibration suppressor system.
Vibration suppression is often utilized to null vibrations associated with a rotating system. Such vibrations, when left unattenuated may lead to crew and structural fatigue and premature failure of system components. The vibrations may also be transmitted through adjacent support structure to other areas and systems remote from the vibration source. Consequently, it may be desirable to suppress these vibrations proximal the vibration source.
One application which exemplifies vibration isolation/absorption is the main rotor system of a rotary-wing aircraft. Typically, the main rotor system includes a hub system which drives a plurality of rotor blades subject to a variety of aerodynamic and gyroscopic loads. For example, as each rotor blade advances or retreats relative to the freestream airflow, each rotor blade experiences a rise and fall of in-plane aerodynamic drag. Furthermore, as the tip of each rotor blade advances with each revolution of the rotor system, the relative velocity at the blade tip may approach supersonic Mach numbers. As such, variations may occur at various coefficients which define blade performance (e.g., moment, lift and drag coefficients). Moreover, gyroscopic and Coriolis forces are generated which may cause the blades to “lead” or “lag.” These effects, as well as others, generate vibrations, which, if not suppressed, are transmitted to the airframe, typically through the main rotor gearbox mount structure.
Various vibration suppressor systems have been devised to suppress vibrations. Mast-mounted vibration isolators suppress or isolate in-plane vibrations at a location proximal to the source. Transmission, cabin or cockpit absorbers reduce vibrations at a location remote from the source.
Mast-mounted vibration isolators having a plurality of resilient arms (i.e., springs) extend in a spaced-apart spiral pattern between a hub attachment fitting and a ring-shaped inertial mass. Several pairs of spiral springs are mounted to and equiangularly arranged with respect to both the hub attachment fitting and the inertial mass so as to produce substantially symmetric spring stiffness in an in-plane direction. The spring-mass system, i.e., spiral springs in combination with the ring-shaped mass, is tuned in the non-rotating system to a frequency equal to N* rotor RPM (e.g., 4P for a four-bladed rotor) at normal operating speed, so that in the rotating system the spring mass system will respond to both N+1 and N−1 frequency vibrations (i.e., 3P and 5P for a four-bladed rotor). N is the number of rotor blades.
While the spiral spring arrangement produces a relatively small width dimension (i.e., the spiraling of the springs increases the effective spring rate), the height dimension of each vibration isolator is increased to react out-of-plane loads via upper and lower pairs of spiral springs. This increased profile dimension increases the profile area, and consequently the profile drag produced by the isolator. The spiral springs must also be manufactured to relatively precise tolerances to obtain the relatively exact spring rates necessary for efficient operation. As such, manufacturing costs may be significant. Additionally, the weight of this device is very high, thus reducing the useful payload of the helicopter. Furthermore, these vibration isolators are passive devices which are tuned to a predetermined in-plane frequency and cannot be adjusted in-flight to isolate in-plane loads which may vary in frequency depending upon flight regime.
Yet another general configuration of a mast-mounted vibration isolator is referred to as a “bifilar.” Bifilars include a hub attachment fitting connected to and driven by the rotorshaft with a plurality of radial arms which project outwardly from the fitting with a mass coupled to the end of each arm via a rolling pin arrangement. A pin rolls within a cycloidally-shaped bushing to permit edgewise motion of each mass relative to its respective arm. The geometry of the pin arrangement in combination with the centrifugal forces acting on the mass (imposed by rotation of the bifilar) results in an edgewise anti-vibration force at a 4 per revolution frequency which is out-of-phase with the large 4 per revolution (“4P”) in-plane vibrations of the rotor hub for a 4 bladed rotor system. The frequency of 4P is the frequency as observed in a nonrotating reference system such as the airframe.
Pairs of opposed masses act in unison to produce forces which counteract forces active on the rotor hub. For the masses to produce the necessary shear forces to react the in-plane vibratory loads of the rotor system, counteracting bending moments are also produced. These force couples may impose relatively large edgewise bending loads in the radial arms, and consequently, the geometry thereof must produce the necessary stiffness (EI) at the root end of the arms. As such, these increased stiffness requirements result in relatively large and heavy bifilar arms.
While the bifilar system has proven effective and reliable, the weight of the system, nearly 210 lbs for one typical system, may be detrimental to the overall lifting capacity of the aircraft. Furthermore, the pin mount for coupling each mass to the respective radial arm may require periodic removal and replacement, which may increase the Direct Maintenance Costs (DMC) of aircraft operations.